Vacuum control of clutch capacity



Feb. 9, 1943. .J. DOLZA VACUUM CONTROL OF CLUTCH CAPACITY '7 Sheets-Sheet 1 Original Filed Oct. 18, 1957 Wk l.

Feb. 9, 1943. J. DOLZA 2,310,518

VACUUM CON'QROL OF CLUTCH CAPACITY Original Filed Oct. 18, 1937 7 Sheets-Sheet 2 mm I Feb. 9, 1943. J. DOLZA 2,310,518

VACUUM CONTROL OF CLUTCH CAPACITY Original Filed Oct. 18, 1937 '7 Sheets-Sheet 5 4" COM J Z w y Feb. 9, 1943. J. DOLZA 2,310,518

VACUUM CONTROL OF CLUTCH CAPACIT Y Original Filed on, 18, 1957 v Sheets-Sheet 4 M w mg niam N Y W n NM nnuv Wu m ll-N Feb. 9, 1943. J. DOLZA 2,310,518

' VACUUM CONTROL OF CLUTCH CAPACITY Original Filed Oct. 18, 1937 7 Sheets-Sheet 5 s I a! g f 11111111114. I

Feb. 9, 1943. J, DOLZA 2,310,518 v VACUUM CONTROL OF CLUTCH CAPACITY Original Filed Oct. 18, 1937 7 Sheets-Sheet 6 Feb. 9, 19' J. DOLZA VACUUM CONTROL OF CLUTCH CAPACITY Original Filed Oct. 18, 1937 7 Sheets-Sheet 7 ENG/NE SPEED 32M Patented Feb. 1943 VACUUM CONTROL OF CLUTCH CAPACITY John Dolza, Flint, Mich., asslgnor to General Motors Corporation, Detroit, Mich., a corporation of Delaware Original application October is, 1937, Serial No.

1940, Serial No. 333,230

Claims.

The invention relates to the control of variable speed ratio in transmissions driven by engines and connected to a varying load, more particularly to pr sent day motor vehicles which are driven by internal combustion engines, and are equipped with step ratio gears requiring transfer of torque from one of a group of driving paths to another, as shown in my application S. N. 169,535, filed October 18, 1937, now Patent No. 2,282,949, dated May 12, 1942, of which the present application is a division.

It relates further to mechanism for automatically selecting speed.ratio, and more particularly to forms of transmissions in which the transition interval occurs without full release of torque, or with torque overlap.

An objeect of my invention is the establishing of a shift sequence in which exceptionally smooth transfer of torque is provided by the coordination of fluid pressure ratio shifting means with regulatory means subject to a measure of torque or to torque demand.

A further object is the providing of regulatory means instantly available and operative at the will of the operator to modify the establishing of drive in a given shift interval according to the existing degree of engine torque.

An additional object is the provision of means subject to the degree of engine intake manifold vacuum applied to the controls regulating the engagement of drive during the shift interval,'

whereby predetermined forces acting on fluid pressure valving are correlated with the driving conditions.

Other objects and advantages to be derived from the use of the invention herein disclosed reside in the interrelation and methods of operation of the parts described, and will become apparent upon inspection of the following specification when read with reference to the accompanying drawings wherein the preferred embodiments of the invention are illustrated.

:It is expressly understood, however, that the drawings are for the purpose of illustration only, and are not to be taken as a definition of the limits of the invention, reference hereunder being noted for this purpose to the appended claims.

In the drawings:

Figure 1 is a stereographic sectioned projection of a transmission installation in a motor vehicle, showing the sequence arrangement of units from engine to load shaft diagonally from left to right.

Figure 2 is an enlarged view of the forward unit of Figure 1 shown in sectioned projection.

Divided and this application May 3,

Figure 3 is a similar enlarged view of the rear unit of Figure 1.

Figure 4 is a schematic view of the control and actuating mechanism of Figures 1, 2 and 3, providing detail of the valving controls and associated parts.

Figure 5 is an enlarged view of the servo and lubrication pump arrangement of Figure 1.

Figure 6 is an elevation section view of the mounting for the valve controls of Figure 4, and shows the external connections linking with the driver and governor connected control members of Figure 1.

Figure 7 shows the driver selection means attached to the steering column of the vehicle, in

view of the transmission assembly of my example discloses a first variable speed unit adjacent the engine connected main clutch shaft 5, having output shaft 8 carrying splined gear I9 engageable with reverse idler gear I8, or with jaws 1' of gear I integral with shaft 5, and constantly meshed with countershaft gear I6 rotatable with gear I1 driving reverse idler gear I0. Countershaft body 20 rotates a first gear I13 of the servo pump drive assembly, hereinafter designated by letter P, and shaft 8 drives a second gear I14 of the servo pump drive assembly, to be described later. Body 20 rotates on I5.

Gear I9 is shifted axially by fork I00 attached to slider I04 of rod IN, and rocker I02 moving in slide I05, the shaft I03 of rocker I02 projecting externally from the casing 2. The above head gearset assembly constitutes a forward-neutralreverse shifting unit, hereinafter designated by letter A.

. As shown in detail in Figure 2, shaft 8 carries integral drum II annually toothed internally at I2 to form the input driving member for the assembly comprising planet carrier 22 aflixed to shaft 2|, reaction sun gear 25 and sleeve 20, and

drum assembly 28-20, the planet gears 24 meshing with annulus l2 and sun gear 25.

Splined clutch hub 34, riveted to carrier 22. carries clutch plate 33 mating with plate 36 held to rotate with drum 28 and presser plate I4. Springs 88 tend to disengage plates 33 and 36. and pistons I2 sliding in cylindrical spaces II in drum portion 29 are arranged to press centering spindle I3 and presser plate I4 to overcome springs 88 and build up capacity in the clutch formed by plates 33 and 36. Fluid pressure is fed through external pipe gland D sage 281 and drilled passage I9 to load plates 33 and 36 to sustain direct drive in the unit constituted by the described planet gearing assembly, hereinafter designated by letter B.

The piston pins I3 transmit loading effort to presser plate I4, plates 36 and 33, the reaction being supported by the intumed flange of drum 28.

Brake 80 encircles drum 28, as shown in detail in Figure 2, to hold drum 28 and reaction sun gear 25 against rotation, whereby reduction gear drive through the gearing is made possible.

When clutch 33-36 is not driving, brake band 80 supported in the transmission casing as shown in Figure 1, is applied, holding drum 2829,- sleeve 26 and reaction sun gear 25 from rotation. With power applied from the engine to shaft 8, and annulus gear I2, the planet gears 24 rotate on their spindles, and cause planet carrier 22 to rotate at a reduced speed, imparted to shaft 2 I.

Shaft 2I extends through unit B, to serve as the power input member for unit C shown in Figure 3 and having two driving sun gears 31-38 aflixed thereto. Forward planet gears 44 meshing with sun gear 38 are rotatably mounted in carrier 54 whose drum extension 52 terminates in internally toothed annulus 5|. Annulus gear 42 meshing with planets 44, is integral with plate 39' fixed to drum 39. The final output shaft 50 is integral with planet carrier 45, rotatably supporting planet gears 43 meshing with sun gear 31 and annulus 5I. Clutch hub 59 splined to shaft 2| is splined externally to support clutch plate 60 mating with plate 55 carried by studs in web 39 and end wall 56 of drum 39, releasing springs 88' as shown in Fig. 3, and similar to springs 88 of Figure 2 serving the same function as springs 88 of unit B, named in this specification as the front unit. Pistons I6 may press on presser plate I8 mounted to rotate with drum 39, through rods 11, the pistons occupying cylinders I5 cut in drum web 56, aflixed to section 33 and web 39'. Actuation of brake 90 of unit C is through thrust rod I90 pivoted thereto at 90a, actieg4 on by rocker I93 pivoted to the casing at The clutches in the drawings herewith are shown schematically as of single plate type, but the invention is applicable by those skilled in the art to multiple plate constructions such as described in S. N. 45,184, filed October 16, 1985, to E. A. Thompson, now matured as U. S. Patent 2,193,304, issued March 12, 1940. The invention herewith is directed to controls for transmission clutches and not to clutches per se.

Fluid pressure is fed to pipe 219' to gland 288, and through passages 281' and I9 to cylinders I5, for loading plates 55-60, to establish direct drive in unit C, known in this specification as the rear unit.

Brake 90 as shown in Figure 3 encircles drum 39, and may look it, and reaction annulus gear 42 against rotation, when reduction drive in unit C is desired, at which time clutch 5560 is released.

Alternate operation of the actuating elements for the forward speed ratio drive of units B and 0 is obtained from the following tabular pattern:

In Figure 4 the controls for obtaining the above shift pattern are shown schematically, as comprising two valves; the first, I50, occupying one of two positions, for admitting servo line pressure from servo main 238 and port 266 to port 26'! and line 218 for actuating unit B for direct drive; or alternately for cutting off servo main 238 and releasing fluid pressure from line 218 to exhaust port 268, for establishing drive through the gearing.

Likewise valve I68 occupies one of two positions for controlling transmission unit C; the first position connecting servo line 213 and port 263 to port 262 and line 219 for establishing di-- rect drive; and alternately cutting ofi line 213 while opening pipe 219 to exhaust through port 260 for establishing drive through the gearing.

Pump assembly P, shown in Figure l, operates continuously as long as either of shafts 5 or 8 have rotation, as will be understood by inspection of Figure 5, which shows the parts in transverse section. The main outlet from the pump is con- 5 trolled by regulator valve 200, which also controls lubrication pressure to the force feed oiling system, not essential to the invention of this specification. Servo main 238 leads from the outlet pressure space of the pump assembly and from valve 200to ports 266 and 263 of valves I50 and I68 respectively, and also through line 330' to port 265 of compensator valve 320 of Figure 4.

For convenience in assembly and manufacture, valves I50, I68 and 320 are mounted in a common valve body 210 adjacent to or integral with transmission casing 2.

As shown in Figure 4, the shift pattern of valves I50 and I68 with respect to the aforementioned shift-actuation table, is as follows:

It will be further understood that when the clutches are engaged in eitheror both units, the corresponding brakes for the respective groups are disengaged by the same fluid pressure which applies the clutches.

The regulator valve 200 of Figure 5 has pressure inlet ports I95 and I36 subject to lift pressure against spring 20I, in ported space 202. When a given pressure head is reached, the lower boss of the valve is raised beyond port 203 leading to check valve 2I9 and servo main 238. At this time leakage through slots 208-2I4 permits flow of oil to port 204 and lubrication line 220. With a further increase in pressure the lower boss will open port 205 and line I91 to set up a balancing pressure on the upper face of the lower boss, whereafter a pressure balance lactween adjustable; spring 2M and thepump pressure is maintaindat a fairly constant pressure point, in that overpressure will blow ofl the surplus through exhaust port 206, and relief valve 2". The overpressure relief occurs when the lower edge of the upper boss permits flow to port 2". The primary purpose of this valve is to afford regulation to pump line servo pressure, and the secondary purpose is to move rapidly to open servo line position with the first incremental rotation of the pump. Other forms of regulating valves may be used without departing from the principles of my invention, the requirement herein being that uniform servo line pressure be sustained in servo main 238. Screw 2I5 adjusts tension of spring 20I.

As shown in Figure 6, valveI68 for unit B is positioned by pivoted bellcrank I60, movable clockwise for shifting the valve to up position, and counterclockwise for shifting it to down" position. Pin I59 of member I60 intersects slot I55 of camplate I26 pivoted adjacent to the pivot I58 of member I60, so that rocking of I26 counterclockwise may cause valve I68 to move to the down position, according to the radial distances of points on cam slot I55 from the center of shaft I52 on which-t26 may turn. External lever II of shaft I52 is connected to manual shifter elements 3I0, I09, 308, 306, 305 and hand lever 30I movable over speed ratio sector positions as indicated on the sector plate of Figure '7. Line 220 of Figures 1 and 5 transmits lubricant to the transmission unit gears and bearings. Leakage passes 288 and 2I4 provide initial flow of oil in the motion of valve 200. Inlet I91 feeds to port 205. Pipe I9I is the suction inlet for pump P, leading to suction space I10. Idler pump gears I82 and I84 mesh with rotor gears HI and I80. Sleeve I18 receives drive from gear meshing with gear I13. Check valve 2| 9 prevents too rapid flow of lubricating oil in line 238.

The inner end of shaft I52 carries afixed lever M5 pressing on spring 4I6, which in turn may press on piece 4 I I pivoted to plate I26, and thereby transmit rotational force through spring M6 to camplate I26, rocker arm I 60 and valve I68. Stop pins on plate I26 prevent departure of levers M5 and 4I1 from the margins of camplate The hand control lever 30I may therefore shift valve I68 to a direct or to a gear drive position for the rear unit C.

In Figure 4 valve I50 receives servo pump pressure from line 238 at port 266. Valve I50 is toggle operated as shown, through toggle amrs I38-I39 pivoted at I31, held by spring I48, and operated by rod II3, but biased by spring I4I toward the active right-hand position. With the valve'positioned as shown in dashed lines in Figure 4, pump pressure from 266 may be applied through port 261 to line 218, to actuate the elements of the brake cylinder 282 and clutch parts connected to passage 19. In the alternative full line position, valve I50 connects port 261 to the exhaust port 268, and servo port 266 is cut oil.

Valve I50 is moved by automatic and manual means, as shown in Figure 6.

Automatic shifter rod H3 is pivoted at its forward end to equalizer bar I I I, pivoted to governor rod H0 at II2. Idler lever I35 pivoted at I35 to valve body 210 is pivoted to rod IIO to limit the motion. Extension 420 of I35 is arranged to intersect the movement of lever 4l1 pivoted on camplate I26.

Cross-shaft 2 of Figures 1 and 5 is driven by gear I15 meshing with gear I14 of shaft 8. Governor weights 25I swing outward with rotation of flange 244 of shaft 24I, the weight arms 260 camming sleeve 233 connected to external fork 350 against the action of governor springs 252- 254, thereby rotating shaft 35I attached to the fork, and rocking lever 352 attached to the shaft and pivoted to rod H0 at 353. The governor assembly will hereinafter be noted by letter G.

It will be seen that through the described linkage, as in Figure l, the variations in speed of shaft ,8, connected to the vehicle engine when the drive is forward, will be transmitted to equalizer bar II I. The accelerator pedal 303 for the engine is connected to lever I32 affixed to shaft I20 of Figure 6 mounted to turn in the casing 2, through rod 355, lever 359, shaft 358, lever 364, and rod 36I. The effective length of rod 36I may be adjustable by any commonly known means, as required for effective operation. The inner portion of shaft I20 carries affixed lever I3 I. Lost motion lever I I6 is pivoted to the casing, and arranged to swing in a path intersecting the motion of lever I3 I. Lever I I6 is drilled at its pivot point M I and connecting passage 4I2 leads to cylindrical space M4, in which slides projecting piston 4I3 engaging the swinging end of lever I3 I.

Pressure line 4 I 0 joins pivot 4| I to pressure lead 219 of the rear unit, as in Figure 4, so that whenever the rear unit is in direct drive, piston 4I3 couples to lever I3I by the volume of fluid existing behind the piston in cylinder 4 I 4. Pin I I5 of lever I I6 may engage equalizer bar III at notch I I4.

Movement of the pedal 303 for advancingthe engine throttle in the normal way through lever 359' and throttle rod 363, will at the same time exert a force through the linkage just described, upon pin II5, tending to cancel or oppose the effect of governor force from rod IIO, by shifting the fulcrum point farther to the left, thereby requiring a higher governor speed for a permitted shift of valve I50 to direct drive position for the front unit B. The throttle pedal connected ratio control linkage above described will hereinafter be designated by letter T.

At the lower portion of Figure 2 is a schematic view of the assembly of control and serve devices constituting the exemplary system for my invention in which brake cylinder 282 is a housing for brake piston 28I slidable on rod 280, piston 285 attached to rod 280, springs 81, 81a, 81b; abutment 300 fixed to retainer 283, and sliding abutment 286 and 288. This assembly is for operating the brake 80 of the front unit B.

The thrust of springs 81, 81a is exerted on piston 28I, and through a shoulder of rod 280, on rocker 393 pivoted to the casing 2 at 394, whereby notch 392 of rocker 393 and thrust rod 390 pivoted to brake member 80 at 80a may apply brake 80 to drum 28 of the front unit. Pressure of line 218 may relieve the spring force on brake 80 by moving piston 28I to the right in the cylinder 282, while at the same time is acting through pipe 218 to load clutch plates 33 and 38 through presser plate'14 and pistons 16 of the front unit B.

The spring forces against which the servo line pressure from 218 is required to work are of predetermined magnitude, so that the resistances to 21 8. It will be seen that if means are provided to apply pressure in line 211 to piston 285. the resistance of the spring system to pressure from line 218 acting on piston 28l will be lessened and varied with the variation of pressure in pipe 211.

Figures 2 and 4 show the construction of the plunger 323 operated by movement of the engine throttle pedal 303 of Figure 1 through changes induced in the degree of engine vacuum. Difierential valve 320 slides in bore 3l9' of valve body 210, the external shell of plunger 323 being bored out internally to fit collar washer 321 slidable on adjacent end of stem 32l of valve 320. Lock ring 340 prevents washer 321 from further motion induced by tension in spring 322.

The uppermost lead 3l8, as in Figure 8, connects the head of stem 3I4 to pressure line 218 of the front unit. The ported passage 265 connects to line 238 (213) receiving net pump output pressure. Ported space 335 is Joined by line 218 to compensator lines 211 and 211', and crossconnected to spac 338' by passage 331. Spring 322 reacts between the lower face of valve 320 and the recessed portion of plunger sleeve 323, guided by stem 32 l of the valve.

With no pump pressure available in either servo line 218 or 219, brakes 80 and 90 are active to establish low speed gear drive in both of the front and rear units B and C by virtue of brake loading springs 81, 81a, 81b, and 91, 91a, 91b.

When the hand control 30I is shifted to move valve I68 to the dashed line position of Figure 4. ports 262 and 263 are no longer connected, and pressure inlet port 263 is shut oil, draining line pressure from cylinder 292 controlling the brake 90 of the rear unit, through exhaust port,260. Pipe Joins pipe 218 to passage 2 (Fig. 6)

The compensator pressure inlet 265 is opened to space 335 and to line 216 feeding both compensator lines 211' and 211 connected to act on compensator pistons 285 and 294 of Figures 2 and 3 of the front and rear units, respectively. Balanced pressure on both ends of boss 399 of valve 320 is permitted by passage 331 and ported space 338'.

Slide 323 may slide in bore 3I9 of valve body 210 and pick up retainer collar 321 to vary the stress of spring 322, whereby the effective aperture between the upper i'ace of boss 339 of valve 320 in space 335 is varied.

Closure of the opening between boss 339 and the seat 355 in space 335 diminishes pressure in line 216 and in connected compensator lines 211' and 211, while conversely, opening of the valve 320 allows the servo pump pressure to be exerted in both of the compensator lines. Exhaust port 336 permits oil to escape from space 355 and line 216.

The hand control of Figures 1 and 7 embodies these members; lever 30l, rod 305, lever 306, clevis 301, rod 308, clevis 309, rod 3I0 and lever II which may rock valve I 68 through the mechanism of Figure 6. Indicator sector plate 302 shows the proper shift positions of lever 30I.

' The distance through which sleeve 323 may move to effect the compensation pressure in lines 211 and 211 is determined by diaphragm 325 of.

Figures 4 and 8 and spring 3I9 attached to sleeve 323, mounted in casing 326 and held therein by clamped shell 328, the casing being attached to valve body 210 by screwed flange 329. Nipple 33I in diaphragm shell 328 is connected to the engine intake manifold 330 through pipe 332 and check valve 333, so as to transmit an index of the engine torque conditions to the sleeve 323. Ring 326 is attached to body 210 and flange 329.

When the engine is stopped, there is no oil pressure acting on the valving, and no vacuum effective on diaphragm 325. The starting of the engine creates a vacuum in line 332 and inside the shell 328, drawing diaphragm 325 down, against the action of spring 3l9 and increasing the compensator port opening of valve 320 between ports 265 and 335. As oil pressure from pump P becomes available in line 235. the pressure may thereupon build up behind pistons 294 and 285 of the brake relieving cylinders 292 and 282.

When control valve I68 for the rear unit is moved to admit pressure from servo line 238 to line 219 in order to disengage brake and applyclutch 55-60, the compensating pressure behind 'piston 294 due to the permitted full flow of compensation pressure by valve 320, allows the clutch pressure in line 19' to act on clutch discs 55-60 at a lower value than if the opposition of brake' springs 91, 91a, 91b, and 910 were fully effective.

Lin 219 transmits servo line pressure from cylinder 292 to clutch cylinders 15 from line 219, byway of gland passage 281' and drilling 19. Spring 284 absorbs the first increment of motion of compensator piston 285. Strap 293 is a retainer for the springs 91.

If, however, the engine throttle be opened before the valve I68 is moved, the reduction in engine vacuum in manifold 330, allows spring 3I9 between shell 328 and diaphragm 325 to shift valve 320 toward closing of port 218 with respect to line 265, which reduces the compensation pressure in lines 211 and 211, thereby establishing a much higher pressure value acting on clutch pistons 16 and discs 55--60, through the direct resistance action of the springs 81 or 91 on the head of the fluid column of the pump. Abutment 214 transmits primary thrust of pin 291 to 910.

The capacity of the clutch 55-60 is controlled by the degree of engaging pressure which sustains it. When, for example, the line pressure of the pump Pis divided between pistons 29I and 294, and springs 91, 91a, 9127 are not fully eflective to establish full pump pressure on clutch pistons 18, the capacity of clutch 55-60 may have a torque value of a: foot pounds, at which time the full range of compensation effect may be present.

Now if compensation pressure be cut oil from piston 294, the full capacity of clutch 55-60 is obtained in a very short time interval, building up a torque value greater than a, or :c+v. Minute changes in between these pressure and capacity levels are therefore predetermined by variations in the degree of engine vacuum, as created by the drivers opening and closing of the engine throttle, and by the driving conditions which likewise aifect the variation of engine vacuum.

Likewise, referring back to Figure 2, compensator pressure in line 211' is exerted on compensator piston 285 of the servo actuator for the front unit.

Conversely, when the car driver advances the engine throttle, the degree of vacuum in manifold 330, pipe 332, and shell 328 diminishes, whereupon spring 3I9 tends to shift diaphragm 325 and valve 320 to closed position, wherein the compensation pressures being no longer sustained, the full pump pressure may build up in lines 218 and 219, according to which of valves I50 or I68 is moved to its direct drive compelling position.

The gradual change of degree of vacuum under these circumstances may be applied to graduation of the opening of compensator .valve 320, in

a series 01' infinitely small steps. which translated .into pressure effect upon clutches 33-33 and 55-30, provides an exceedingly smooth engaging action, so that there are no discernible lurches or Jerks in the engagement of either clutch involved in a speed ratio shift sequence, in either ascending or descending ratio changes.

In Figure 8 the compensator valve 320 of Fi ure 4 is moved by'spring 322, sleeve 323, and arm I22 of shaft I23 mounted in valve body 210. Shaft I23 projects externally from body 210, and is rocked by attached lever I21 pivoted to diaphragm rod 400 attached to diaphragm 325' mounted in casing 326' and shell 328', whose extension carries nipple 33I and afiords a seat for spring 3 I9 bearing against the diaphragm. Rod

400 is a loose fit at port 432 in shell casing 326' sure of the flow from space 338 to piston 234 so that if desired, the release or intake of air in the compartment between diaphragm 325' and casing 326' may be regulated to dampen or to increase the sensitivity of the action or. the vacuum from manifold vacuum lead 332 on the compensator valve 320.

The center of shaft I23 is so taken with respect to the center of valve 320 and plunger 323, that rapid initial motion followed by slower motion in the stroke of plunger 323 with respect to variation of vacuum upon diaphragm 325', may be obtained.

If converse action is required, the center of 523 is located above the point of contact of I22 with 323. This expedient makes it possible to extend the use of my invention to a wide range of power plant installations wherein torque responsive control conditions. with respect to available power and load to be driven are widely different from those experienced in customary passenger car practice.

The purpose of this control is to proportion the rate of clutching engagement or capacity change to the degree of torque demand existing at the interval when the clutch is selected to drive. I

If the torque demand is falling off during the clutch engagement interval, the full pressure available from the servo pump might become effective before the relative rotations andinertias of the parts had diminished in accordance,

extent such that the end of stem 32I meets the inner end wall of plunger 323, and positive clomay occur.

The chart of Figure 9 provides a series of characteristic internal combustion engine torque curves for varying conditions of engine speed, throttle opening and degree of engine intake manifold vacuum.

The dashed lines represent the engine torques at different engine speeds at given throttle openings, as follows:

Curves Throttle a-a Partial. b-b uarter.

e-c all. d-d 3-,quarters. e- Full.

It will be noted that these curvesfall off. rather sharply with increased speed.

The full line curves represent the engin torques at different engine speeds at given degrees of engine intake manifold vacuum, diminishing with torque from curves I to IV. These curves are relatively flat and yield a close approximation to the torque values.

This set of curves demonstrates the utility of my invention in providing a uniformly varying change of degree of vacuum with torque, from which resultant force, the coordinate applications of Figures 4 and 8 are obtained. The close agreement of the vacuum values with the engine torque makes the application of my invention a mac-- whereupon the clutch connected elements would be joined quickly at full loading capacity. This would administer a shock to the transmission system, and cause a lurch of the vehicle, since the absorption period for momentary differences in torque at the clutch discs is very short.

At this point, the vacuum responsive valve 320 interposes the diminishing torque demand factor, and the rise in engine vacuum moves diaphragm 325 to compel compensator valve 320 to shift toward an opening position, wherein increasing compensation values operating on compensating pistons 285 and 294 are obtained.

The first increment of accelerator pedal motion from idling causes the degree of vacuum in manifold 330, pipe 332 and shell 328 to decrease, permitting spring M9 to load light spring 322, so as to oppose the force of fluid pressure acting on the upper face of valve 320. This results in a graduating of the orifice between the lower lip of port 335 and the upper face of valve 320, restricting the pressure flow from space 334 to outlets 2'I62'l'|, available to create pressure on compensating piston 234 in the cylinder space between abutment wall 295 in cylinder 292.

At full accelerator pedal position the existing degree of vacuum is opposed by spring 3 I9 to an tical utility, and yields an instant response when operating over highways in territories of con-- siderable gradient, so that the ratio shifting clutches are provided, with a torque capacity in close accordance with the actual need.

The arrangement of gearing, clutching and. braking shown herewith, provides a ratio shift sequence in which during forward drive, the transitions from one ratio to another always occur with a given mean torque value between input and output. In other words, at no time in forward speed drive is there a neutral, or nodrive condition while ratio is being shifted. The overlapping of torque among the shifter elements such as the clutches and brakes used in the illustration provides an extremely smooth method of passing through the transmission shift inter vals, but practice has disclosed that supplementary means to adjust the torque capacity of the clutches about to take up drive, in accordancewith the existing torque demand is required in order to absorb the existing inertias of the rotating parts, and to avoid abuse of the friction members which have to carry the torque.

It has been shown in the clutch art that regulation ofthe rate of clutching engagement by varying the position of the engine throttle provides a certain control over the smoothness of clutching action, but this, however, has been applied to friction clutches pro-biased for driving,

the control being upon the permitted rate of engagement. In my invention the ratio shifting clutches are pre-biased for disengagement, and I superimpose my method upon asystem in which automatic and manual controls not only initiate the clutching action but also apply a given rate of pressure build-up to establish a predetermined torque capacity level for the clutch or clutches involved.

The end point of the clutching action on the rising pressure phase of the clutch control cycle is thereby determined by the torque demand of the driver as expressed by the movement of the engine accelerator pedal, and further modified by the driving conditions, in that the degree of engine intake manifold vacuum varies with both characteristics.

The invention further equilibrates the forces derived from predetermined throttle opening and driving conditions against a spring, for example, item 3i9 of Figure 4, of given rate, and likewise against a predetermined fluid pressure force'acting in opposition upon valve 320, so that for the range of movement of the compensating control, an extremely flne adjustment of the fluid pressures acting on the clutches 3336 and 55-60 is obtained. This method is in no way sensitive to wear of the clutches, in that pressure of the constantly operating pump is always available to take up slack due to wear, by creating a volume increment of corresponding value.

Variations in viscosity of the fluid used in the system can only affect the friction of valve 320 in bore 3i9, the fluctuations of viscosity being of magnitudes of less effect that the calculated forces available from vacuum line 332, diaphragm- 325, spring 319, spring 322, and pressure lines 265 and 3H3.

Among the structures to which the present invention is applicable are those described and shown in the publication known in the automotive trade as Automotive Industries, for May 29, 1937, on pages 806 to 809 and 823.

Thus I have provided a novel driving mechanism and control which is effective to achieve the objects above enumerated, which possesses advantages in manufacture and servicing, and which is adaptable to a wide variety of use and applications. For example the system shown is useful for driving rail cars, aircraft propellers, superchargers for induced air combustion draft, tractors, farm and excavating machinery. Wherever a drive between a variable speed and power engine and a variable load is required the demonstration of my invention is applicable.

It is herewith acknowledged that changes from my construction and arrangement of parts will suggest themselves to those skilled in the art, but it is,understood that such means are within the scope of the invention herein disclosed and as defined in the appended claims.

I claim:

1. In automatic controls for driving mechanisms, in combination, an engine, a variable speed control for said engine movable into various positions representing the torque demand of the operator, a clutch driven by said engine, auxiliary power means arranged to actuate said clutch, regulating means for said clutch connected to said power means effective to predetermine the torque capacity of said clutch, and a device responsive to changes in engine vacuum and subject to the variations of said speed control and acting on regulating means such that for increased operator torque demand the torque capacity of said clutch is increased.

2. In variable controls for driving mechanisms, in combination, an engine, a variable speed control for said engine, a clutch driven by said engine, control means for said clutch, a device connected to said means arranged to vary the torque capacity of the clutch, and means responsive to changes in engine vacuum subject to the variations of said speed control acting on said clutch control means such that for advanced positions of the control the said torque capacity is increased, and for retarded positions of the control, the said torque capacity is diminished.

3. An engine, a speed control pedal for said engine, a loadshaft, a friction clutch for transmitting the power of said engine to said shaft comprising mating driving elements, engaging and disengaging control means for said clutch including control mechanism effective to establish selective driving pressures at which said clutch provides initial drive, clutch torque graduating control means operative to regulate the driving pressures between said elements, and engine vacuum responsive means efiective upon said graduating control means such that the selected driving pressures on said elements vary with increases or decreases in engine speed.

4. In transmission clutch control means, in combination, an internal combustion engine, a vacuum force generating device responsive to the power conditions of the engine, driving and driven shafts, a friction clutch adapted to transmit drive therebetween, loading means effective to establish drive between input and output members of said clutch, a differential valve active on said means arranged to select the loading pressures on said clutch over a predetermined range of pressures at which the clutch transmits drive, and control means responsive to variations in the degree of engine vacuum acting upon said device arranged to position said valve at the will of the operator.

5. In automatically controlled variable speed transmissions for motor vehicles, in combination, an internal combustion engine, a vacuum manifold connected to said engine, a clutching device adapted to transmit the power of said engine composed of two groups including groups of mating friction elements each group of which is arranged to rotate as a unit, means for engaging the device subject to both loading and unloading forces, control mechanism arranged to engage and disengage said device through said means according to a predetermined operating cycle, and auxiliary controls responsive to variations in the degree of vacuum established in said mamfold operative to vary the rate -of engagement according to the torque demand of the vehicle driver.

6. In automatically controlled variable speed transmissions for motor vehicles, in combination, an internal combustion engine, a vacuum generating device responding to the power conditions of the engine, clutch mechanism adapted to transmit the power of said engine, a servo pump, a servo pump line, and a clutch actuator for said mechanism operating from said line through a valve, a pressure line connecting said actuator and said valve, a fluid pressure accumulator connected to said pressure line, an auxiliary accumulator connected to said servo pump line through a second valve, and control means subject to the vacuum acting on said device operative upon said second valve effective to diminish or increase the rate of pressure in said pressure line and actuator at the will of the operator.

7. In automatic variable speed power trans.

mission gearing, in combination, an engine, a vacuum force generating device responding to the power conditions of said engine, clutching mechanism adapted to transmit the power of said engine, including a clutch pressure regulating system comprising a cylinder, pistons in said cylinder, a piston rod moved by said pistons. a fluid pressure conduit means connected to said cylinder, springs arranged .to be stressed by fluid pressure acting against said pistons, clutch actuation means connected to said fluid pressure conduit, a clutch embodied in said mechanism and actuated by said means, and fluid pressure control mechanism subject to changes in the degree of engine vacuum and connected to said fluid pressure conduit operative to initiate an increase of pressure in said system for establishing a predetermined torque capacity in said clutch.

8. In automatically operated gear mechanisms, in combination, an engine, a speed control pedal for said engine, a clutch connecting said engine through gearing to a load shaft, fluid pressure motor means arranged to sustain drive in said clutch, supply means for said motor means, control means operated by said pedal for the application of fluid pressure from said supply means to said fluid pressure motor means, valve means to vary the engagement and driving pressure to ans positions representing the torque demand of the operator, a clutch driven by said engine, servo motor power means arranged to actuate said clutch, regulating means for said clutch operative upon said servo motor power means effective to predetermine the torque capacity of said clutch, and a device responsive to variations in .the engine power condition subject to the variations of said speed control and acting on said regulating means such that for increased operator torque demand the torque capacity of said clutch is increased.

10. In variable controls for driving mechanisms, in combination, an engine, a variable speed control for said engine, a clutch driven by said engine, actuation means for opening and closing said clutch, control means for said actuation means, a device connected to said control means arranged to vary the torque capacity of the clutch, and means responsive to variations in the engine power condition subject to the variations of said speed control and acting on said clutch control means such that for advanced positions of the variable speed control the said torque capacity is increased, and for retarded positions of the variable speed control, the said torque capacity is diminished.

JOHN DOLZA. 

